JPH03281986A - High-pressure device - Google Patents

Abstract

PURPOSE: To improve delivery efficiency by installing two intermediate pressure pistons, and periodically driving two high-pressure pistons by these intermediate pressure pistons alternately, while installing a film member or a fluid separate means to separate two fluids in both inward and outward chambers from each other. CONSTITUTION: This high-pressure aggregate which is provided with an inlet opening 487 to be connected to an outlet opening 488 of an intermediate pressure generating unit consisting of a pump unit 19 with a power source 401, has a control unit 17 to be rotatively driven by a fluid motor 97 being driven by a return flow via two reduction gears 467 smf 466, and this control unit feeds two cylinders alternately with an intermediate pressure fluid from an opening 91 and two fluid passages 472 and 473, periodically driving two pistons 8 and 9 alternately in these cylinders. These intermediate pressure pistons 8 and 9 drive both small diametral high pressure pistons 5 and 6. Each of these high pressure pistons 5 and 6 is installed in a housing 464, imbibing the fluid through both inlet/out parts 64 and 65, and they discharge it as a high pressure fluid.

Description

DETAILED DESCRIPTION OF THE INVENTION [Technical Field of the Invention] The present invention relates to
The present invention relates to high-pressure devices, such as high-pressure pumps, capable of generating or discharging a high-pressure fluid flow, such as a high-pressure fluid flow extending over 1000 yen.

[Conventional technology]

In fountain flow cutting and other high-pressure fluid utilization techniques, high pressure is generated by a so-called "shaft booster," which uses a relatively large-diameter piston that is hydraulically driven, which in turn drives a relatively small-diameter high-pressure generating piston. It is widely practiced to obtain A high pressure (generating) piston acts on water and discharges it outside the device. The problem is that the piston seals are difficult to construct and suffer from large efficiency losses due to high levels of friction.

In the past several years, the present inventor (and applicant) has obtained numerous United States patents regarding high-pressure water pump devices configured to sequentially and periodically operate a plurality of pistons. but,
These patented pump devices are expensive to manufacture, and they still have some problems with uniformity of fluid flow. Furthermore, they also have the problem of requiring a relatively heavy speed reduction device between the drive electric motor and the pump device.

Accordingly, there is a need for further improvements in conventional high pressure equipment, and the present invention can achieve at least some of the required improvements.

As mentioned above, the high-pressure device for generating high pressure of about 4,000 atmospheres is the booster, which uses hydraulic oil to drive a piston with a relatively large diameter. drives a relatively small piston (high-pressure piston) to discharge water at high pressure. It is difficult to seal the high-pressure piston of the high-pressure water pump device, and the service life of the seal is relatively short. Furthermore, every cycle of operation of the booster (high-pressure device), there will inevitably be times when the booster (high-pressure device) does not send or discharge water.

Therefore, a normal booster is provided with a pressure accumulator in the fluid flow path for such a time of non-operation. This pressure accumulator is a device or means for delivering water at high pressure when the booster is not delivering water, ie, when it is inactive.

The above-mentioned problems with boosters are addressed in US patent No. 4.690.6 owned by the inventor,
No. 20, No. 4.701.113, No. 4.745,
No. 846, No. 4,799,654, No. 4.824
．． No. 338, No. 4, No. 822.255 and No. 4.
904.167 is solved by a continuous water pumping device. The pump device disclosed in the above-mentioned patent continues to pump continuously during each rotation of its drive shaft, thereby eliminating periods of inactivity when the booster is not pumping. However, these devices are relatively complex in construction, expensive, and still require sometimes difficult sealing. Therefore, it is necessary to simplify the structure of these devices, increase their reliability, extend their useful life, and reduce their manufacturing costs.

[Object of the Invention, Technical Background and Overview of Technical Content] Accordingly, the main objects of the present invention are to improve known high-voltage devices to increase their reliability, extend their service life, improve their sealing configuration, and The objective is to improve fluid flow uniformity and further reduce manufacturing and maintenance costs.

As mentioned at the beginning of this section, the present invention relates to a high pressure (generation) device or pump. One feature of the device of the present invention is that
The advantage is that it can be used as a pump device for non-lubricating fluids.

FIG. 10 of the accompanying drawings is a sectional view of a prior art "axial booster." Intermediate pressure fluid is sent alternately into chambers A and B from openings ■ and W, thereby causing intermediate pressure piston K
is periodically driven back and forth from side to side in FIG.

Control of the stroke direction of the piston may be provided by conventional control devices. Pistons d, which have a smaller diameter than the diameter of the piston (or drive piston) K and extend in the axial direction, are provided on both sides of the piston. These small diameter pistons (or high pressure pistons) d and d extend within the high pressure cylinder or chamber and E and perform reciprocating movements within them. The high-pressure cylinders D and E are closed off at their axially outer ends by cylinder heads. The cylinder head is provided with an inlet and an outlet, which are well known and are not shown in FIG. 10.

First, let's assume the following. That is, the stroke of the piston assembly K-d is between A-8 and the change in direction of this stroke occurs within zero (0) time (i.e.,
Assume that no time is spent on this conversion) and that there is no dead space in the piston chamber and around the valve.

A device that satisfies these conditions may exist in reality.
It is a theoretical device. In reality, this means that some time is spent changing direction in the journey, and boosters like this (
This is because there is some dead space in the device.

Chambers A and B are alternately subjected to intermediate or medium pressures P. When one of chambers A and B has a pressure P,
The pressure in the other chamber is 0 (zero). That is, when the pressure in chamber A is zero, the pressure in chamber B is zero, or vice versa.

When the piston stroke in either direction begins, the pressure in the chamber is zero. In the present invention, it is important to recognize that fluids are compressible. For example, water has an intermediate compression coefficient Fctn=0.0000375, and hydraulic oil has an intermediate compression coefficient F,. -0,000054. This means that water is compressed by 0.00375% due to pressure increase of Ikg/c+11, while hydraulic oil is compressed by 1kg.
/c means compression by 0.0054% due to JO pressure increase. These considerations target intermediate or intermediate pressure regions within a wide range of pressures, and ignore the effects of temperature and/or viscosity. The outlet valve of cylinder D is set to a high pressure of, for example, 2000 kg/afl.

The actual volume in cylinder D is d"-7,×stroke S. Since high pressure exists in the discharge passage after the outlet valve, the As long as the pressure is greater than the pressure in the cylinder D, the outlet valve will not open.The high pressure in the discharge flow path after the outlet valve is defined as Pl.

Then, after the start of the piston stroke, the outlet valve opens and discharge of the high-pressure fluid begins after the fluid in the cylinder D is first compressed to the high pressure P6. In other words, a part of the piston stroke is wasted ineffectively for discharge, resulting in a volume loss Vbc according to the following equation (1).

Vdc=d”(W/4)XSXFcwXPh (
1) However, the above is a case where the fluid to be discharged, that is, the fluid in the cylinder D is water. If this fluid is (working) oil, the compression coefficient pc of the oil can be substituted for FCW in the above equation (1).
Insert o.

As a concrete example, (piston) stroke S is 42 m, the diameter of chambers or cylinders B and D is 80 mm, chamber or cylinder E
Considering the case where the diameter d of
42 X O-0000375X 2000= 193
9-62m5+3.

The volume v4 of the cylinder (chamber) before starting the stroke S is ν
4=d2・1・S=28”・1×42・25861.
It is 6ms+3.

By dividing the volume loss VdC by the volume v4, it is possible to calculate the compression loss in each chamber or cylinder as a value of 1/100%, and if the compression loss of the fluid in each chamber is vc, it is calculated by the following formula. It is possible.

Vc = Vac/Va J, 01% (2) Using this percentage value, the loss in stroke length due to fluid compression can be directly determined, and if this loss is S, then Sc = SXVc (3) .

In the case of the above specific example, the loss VC is VC=1939.62
/25861.6=0.075 and the loss SC is 5
c=42 x O, 075-3,15 (mm).

During this 3.15 mm portion of the 42 tmh stroke S, the high pressure Hp is not discharged from the outlet valve, and this loss in the above example is 7.5% (3,15/42
).

However, if cylinder B has a dead space or is in communication with dead space N, the volume of this dead space must also be compressed before the pump device can discharge high-pressure fluid. , this also applies to dead spaces M, G, and H in chambers A, D, and E.

If these dead spaces are included in the calculation, the loss VCt
is as follows.

VCL= (Vd+Vg) X PhX Fce/V
a (4) Also, taking into account dead space,
Stroke loss Set is as shown in the following equation.

5cL=SxvcL (5) In the above specific example, the volume of dead space G is 400
Assuming 0 mm3, the calculation is as follows.

VcL-(28861+4000) x 2000 x
O,0000375/25861.6-0,0866 Therefore, the loss in discharge time due to compression of water in the room is 8.66%.

From the above it will be immediately understood how important it is to reduce dead space as much as possible.

For more complete consideration and calculation, large diameter chambers A and B
It is necessary to consider whether the intermediate pressure fluid in the chamber is oil and the intermediate pressure chamber and the high pressure cylinder have a dead space or are in communication with the dead space. In the above specific example, the intermediate pressure P is 300 kg/c1il, and the dead space of the intermediate pressure chamber is 6C=6000+am'
Assuming that, the following calculation formula is established and the calculation result is obtained.

Vcctt= [(Vn+VJ −Fcti・Ph/V
o] +Vn・Pco・P−/Vo (
6) Substituting the numerical value in the above specific example into this second, VCLL= [(25861,6+4000) X O
,0000375X2000/25861.6] +6
000 x O, 000054x300/25861.
6 = 0.09036 In an actual pump device, the piston must not touch the cover (cylinder), so there is a small space distance that causes loss before the stroke actually starts, and this volume is also calculated by the following formula: must be taken into account as follows.

Vq=(D"-d Lee>(Q is distance WB> (7
) and Vcttt= [(Vo+Vc) ・Fcw-Ph/V
ol + [v, 4+(o"-d")・-,Q], FC
O, P, C D (8) The above equation (8) can be rewritten as follows.

Vctrt= ((Vn+Vc,)・Fcw・Ph +
[VN+(o"-d"), -, mouth/Fco-P-)
Ha. (9) In the following calculation example, to high pressure = 4000
kg/CT1 or intermediate pressure P, = 600 kg/cdl
It is assumed that

ttt=[(25861,6+4000) X O,
0000375X4000+ [6000x (80”
-28”) x 7 xO, 00054x600/25
861.6=0.2028, that is, the discharge time loss Fv is
Fv=100XVcttt·20.28(χ).

Such a catastrophic loss is completely unacceptable in modern technology, and it will be understood from the above description that there is an urgent need to improve the pump device shown in FIG. It is an object of the present invention to provide improvements in this and other respects.

[Examples and effects of the invention]

FIG. 11 of the accompanying drawings is a sectional view showing a part of a high-pressure device related to the invention of the present inventor, which has not yet been published or announced. The device is provided with two intermediate pressure pistons K, which drive high pressure pistons P and P in cyclic alternation. The high pressure piston P pumps fluid into the outer chamber OC. A membrane member or other fluid separation means is provided to separate the fluid in the inner chamber IC and the fluid in the outer chamber OC from each other, the membrane member or other fluid separation means is axially flexible,
It can move into the outer chamber and the inner chamber. It is possible to have a lubricating fluid in the outer chamber OC, while a non-lubricating fluid in the inner chamber IC. These chambers or cylinders 0C1IC may also have a dead space or communicate with a dead space, as in the above case.

From the knowledge obtained from the considerations described above, the following equation can be obtained for the case of the apparatus of FIG.

pcw・[Vco・Pc1n・Ph+(Vo+νN)
・Fco・Ph+D”・, ,Q,PC,,P,]
QO) and VcaH□ [2Voc・pcw−to(νD+VN)
・Ph+D!・1・0-FCo-P, j/v, (11)
By substituting the number 4M (pressure value is Ph=4000kg/ctA and P=600kg/c+ll) in the specific example described above into the above equations 00) and 01), the following results are obtained.

VC,= [25B61.6 X O,0000375
X 4000+ (25861,6+6000)Xo,
00054x4000+80"X -X 4 X0, 0
00054 x 600] /25861.6=0.4
413 and VcdH = [2X 25B61.6 X O,000
0375X 4000+(25861,6+60000
) X O,000054X 4000+80”
・0.5913 Therefore, the discharge time loss is 0.5913.
4413X100・44.13% and 0.5913X
It is 100.59.13%.

Note that the above equation 0ω applies in the case of the device shown in FIG. This applies respectively to the device of FIG. 12 forming inner and outer chambers.

A surprising result arises from the above considerations. That is, by using two sets of pistons that alternately reciprocate, -
Although it may seem that a uniform fluid flow can be obtained at all times, the fact is that the pressure of the discharged fluid is very high, and if the device is configured compactly (small and high performance) (No. 11) and the apparatus shown in FIG. 12),
Dispensing time losses amounting to approximately 59% and therefore fluid flow non-uniformities amounting to approximately 59% result.

Many of the problems of the prior art device shown in FIG. 10 are overcome or ameliorated by the example device shown in cross-section in FIG. 1, which depicts one embodiment of the present invention. In particular, improvements are realized in the sealing arrangement of high pressure (generating) pistons and problems with electrical or electronic application controls are eliminated.

FIG. 1 shows in its upper part an embodiment of the apparatus of the invention, and in its lower part a conventional intermediate pressure bomb apparatus 19. This latter pumping device 19 is driven by a power source (plant) 401 and does not deliver intermediate pressure fluid through outlet opening 488 .

The embodiment of the high pressure system of the invention shown in the upper part of FIG. 1 is a complete system in itself, with the power source 401
and an outlet hole 488, the pump device 19, ie, an intermediate pressure generating unit, can be transported to any location or device where an intermediate pressure generating unit is available. Such locations or equipment are, for example, various manufacturing or workshop rooms, power shovels, ships, etc., in which intermediate pressure generating units such as those described above are usually provided. The embodiment of the invention shown in FIG. 1 has an intermediate pressure inlet opening 487 connected via a fluid line to an available intermediate pressure generating unit outlet opening 488. An inlet opening 487 is provided in flange 486 and is in communication with controller 17 by fluid passageway 408 . This control device 17 is rotationally driven via speed reducers 467 and 466 by a fluid motor 97 driven by the return flow from the high pressure device of this embodiment. The control device 17 alternately directs intermediate pressure fluid through openings 91 and 93 and fluid passages 472 and 473 into the cylinders or chambers 14 and 15 to alternately and periodically drive the pistons 8 and 9 therein to complete the drive stroke. have them do it. These intermediate pressure pistons 8 and 9 drive high pressure pistons 5 and 6 of small diameter compared to their diameter. The fluid in the intermediate pressure chambers 44, 45, 46 and 485 drives one of the piston sets on a downward return stroke, while this piston set performs an upward drive-discharge stroke. A pump 484 is provided, which fills the intermediate pressure chamber with fluid from the fluid flow path 485, but a safety valve is provided in the discharge passage 413 to fill the intermediate pressure chamber 4 with fluid.
4.45.46 and 485 may be configured to maintain precise amounts of fluid and pressure within them. When each piston set performs a return stroke or return movement, fluid flows from cylinders 14 and I5 to fluid passages 472, 473 and control opening 392.
, and is further sent through fluid passages 94.301.302 to a fluid motor 97, which in turn drives the control device 17. Around each opening of the motor 97 that performs rotational movement, there is a sealing ridge or wall 394.393°392.
are formed at the side ends in FIG.
can be provided. Fluid passage 408 also includes space 41
0 to a safety valve 409, which allows fluid to flow from this passage 408, as required, through the safety valve 409 and fluid passages 412.301 to the return flow motor 97. This configuration is important. This is because if for some reason one of the piston sets stops at the bottom of the stroke, without the above configuration, no fluid will be sent to the motor 97, and the entire device may stop working. Because there is.

The high pressure pistons 5 and 6 are located within a housing or body 464 which is fitted with a valve head 489 having a mouth valve and an outlet valve.

Reference numeral 64 designates the inlet/outlet to the high-pressure cylinder or chamber II, while reference numeral 65 designates the inlet and/or outlet to the pressure cylinder or chamber 12. With the bolt, the valve head 4
89 to the housing 464, the seat 490 is for supporting the bolt heads of those bolts.

According to the invention, the high pressure pistons 5 and 6 are such that their parts reciprocate in different cylinders (chambers) containing different fluids. However, preferably each of these high pressure pistons 5 and 6 is of constant diameter over their entire length. High pressure piston 5
and 6 are part of the housing 464 and are disposed within a first seal proximate to the intermediate pressure chamber and hermetically isolated from the intermediate pressure chamber by lubricating oil, in which reciprocating movement is possible. It is designed to do this. At the opposite ends of these first sealing parts, chambers 451 and 452 are formed which are first return/unloading chambers and are leakage fluid collection chambers. Leakage of lubricating fluid from the first seal is collected in these collection chambers 451, 452, and the collected fluid is directed from fluid passage 457 to outlet opening 457°.

An intermediate seal 453,454 is provided at one end of the first liquid collecting chamber 451,452, and a second liquid collecting chamber 453,454 is provided on the opposite side.
It will be appreciated that leakage fluid collection chambers 455, 456 are provided. These second collection chambers 455.456
The leakage fluid from the second seal formed at the opposite end of the collection chamber 455, 456 is collected, for example by a bayonet bushing 459 or 460. The fluid collected in the second collection chamber 455, 456, which is often a non-lubricating fluid such as water, is directed from the fluid passageway 458 to the outlet opening 458'. Although the housing 464 is shown divided by dividing lines 492 and 493, the housing 464 is shown divided by dividing lines 492 and 493.
64, a sealing portion is further provided there, and the first and second liquid collecting chambers 451 and 4
It is also possible to provide a housing part defining a fluid outlet between 55 and/or 452 and 456. Plug-in bushes 459 and 460 are connected to the housing 4
64, which form a high-pressure seal for the high-pressure cylinder parts 11 and 12 and the high-pressure piston 5 and/or 6. The space in the housing 464 into which the plug-in bushing is inserted and where the first seal is formed can be precisely machined to have a precise cylindrical surface, so that it is suitable for the piston 5 or 6. The cylindrical inner surface of the plug-in bushing forming the cylindrical second sealing surface of the cylinder portion 11 or 12 of the high-pressure piston 5
It is possible to form it so that it has exactly the same axis as each of 6 and 6. Therefore, it is possible to fit the high-pressure pistons 5 and 6 into the cylinder parts 11 and 12 and to seal therebetween extremely tightly and accurately. This sealing configuration can be achieved with a high degree of accuracy by making the axis of the piston and the axis of the cylinder part concentric with a high degree of precision, so that the space between the piston 5.6 and the cylinder part 11.12 is
A tight fit with a gap to allow sliding of the piston provides a seal against high pressures. Also in the device parts not shown between the dividing lines 492 and 493, corresponding parts of the respective pistons can be configured to be driven by the same fluid. Therefore, a sealing section similar to that described above can be configured in this device part as well. Liquid collection chamber 451
．． Since the pressure applied to 452° 455, 456 and seals 453, 454 is low or "zero" pressure, seals 453, 454 are made of weak metal that is not likely to cause wear to the outer surface of piston 5.6. A low-pressure seal formed by

The above embodiments of the invention make it possible to obtain the sealing technique described below.

Let Q be the fluid leakage from the first and second sealing parts shown in Figure 1. If we ignore the effects of eccentric movement of the piston and changes in the piston diameter and fluid viscosity due to temperature changes, this is as follows. It can basically be calculated using formula 〇2).

In the above formula, Q: Leakage (C1ll/5ec) η = (Fluid) viscosity (kgs/Bo) P = Positive pressure kg/c+fl) di = (Cylinder) inner diameter (mm) L = Sealing length (mm) δ = Gap Radius (+A of diameter) (Peng) π=3.14 P+, Pg- Pressure at both ends (operating) Oil pressure (ηoff) is, for example, ηoil=0,
00264kg5/nf and that of water (y2
water) is, for example, ηwater=O, OO006k
gs/rrf.

The pressure difference between both ends of the sealed part is 2000 kg/c1a
, gap radius (δ) is 0.005mm, oil seal length is 4
0mm and the water seal length is 60mm, the following calculation can be done (However, if the piston is 28mmφ
).

Oil 10 π 280otl = "z
, 0.00264 ・2000' 4o(0
,005)”=0.1735c4/5ec=10.
41csff/win water Gwater--2000 ・translation (0,005)
'lOπ 12 0.00006 60=5.9cJ/
5ec=305cIl/win In the apparatus shown in Fig. 1,
The piston diameter is 28mm, and this device has a piston diameter of 101/m.
If we send 1n=10.000cJ/win of water,
The total amount of water leaked is 305cIiI/11inX 2 = 6
10c hair/win or 6.19% of the theoretical feed amount (・
610/10.000・0.0619x 100). In contrast, in the sealed piston configuration of the prior art device shown in FIG. 1O, 10 to 20% of the total force is lost due to friction. From this, it follows exactly the construction principle of the device shown in Figure 1, and as mentioned above, the gap radius is 0.006M (or gap diameter: 0.01mm
), it will be understood that the device shown in FIG. 1 is far superior in efficiency compared to the conventional device shown in FIG. 1O.

Leakage is less with oil compared to water. This leakage ratio is 4000kg/c! In the case of high pressure, it becomes a problem. This is because in such a case the leakage would be doubled to approximately 12.38%.

In this case, the apparatus of FIG. 1 and the apparatus of FIG. 10 are approximately equal in efficiency (ignoring the compression of each fluid).

If we double the gap, i.e. make it 0.
01mm or 0.02am in diameter, the first
The device shown is even more disadvantageous. In this case, the leakage is 8
double, and the device of FIG. 1 is inferior in efficiency to the conventional device of FIG. 10 (again, ignoring fluid compression). Therefore, as mentioned above, it is necessary to make the axial center of the piston and the axial center of the cylinder part perfectly coincide, and to accurately reduce the gap between the outer surface of the piston and the inner surface of the cylinder part. It will be appreciated that it is necessary to manufacture the device of FIG. 1 precisely according to the principles of FIG.

If such a high degree of accuracy cannot be achieved, or if such leakage is unacceptable, the device of the present invention, shown in the other attached drawings, may be used instead of the device of FIG. Other embodiments of the device may be employed, such as those utilizing a membrane or other fluid separation means (hereinafter referred to as a separator) between the inner and outer chambers. Such an apparatus is shown in FIG. 11, FIG. 12, FIG. 47, FIG. 65, FIG. 40, FIG. 48, or other figures.

FIG. 1, together with FIG. 2, FIG. 3, FIG. 4, FIG. 5, FIG. 6, FIG. 7, FIG. 8, and FIG. Other embodiments of the present invention are also shown that partially or completely overcome the problem of compressive fluid losses.

A feature of this embodiment is the formation of an auxiliary flow in the fluid flow, which immediately fills the cylinder part with the compressed fluid at the change of direction of the piston stroke, which is then expended for fluid compression as already mentioned. The purpose of this invention is to eliminate the discharge time loss which was the cause of discharge loss and non-uniformity of discharge flow.

For this purpose, a second or auxiliary control section 517 is provided on the back of the main control device or controller 17 of FIG. 1, as shown in the cross-sectional views of FIGS. 2 to 9, for example. This auxiliary controller can be rotated in full alignment with the main controller 17, but with the openings or holes reoriented, as indicated by the reference numerals in FIG. The auxiliary flow can be passed through the inlet opening 1489 to the auxiliary control section 517 and through the fluid passageway 474.475 to the cylinder section 14 or 15. Conversely, it is also possible to cause the auxiliary flow to flow from the cylinder section into the control hole of the auxiliary control section 517 through the feed outlet opening 1493.

It may also be necessary to provide a fluid passageway 301. The above-mentioned control hole of the second or auxiliary flow control section 517 is an opening or fluid passageway 491, 492, 490 or
Other openings or fluid passageways shown in the figures may be used. Practical design of these control holes, not shown in the drawings;
Please ask the inventor for the configuration. As previously mentioned, the auxiliary flow rapidly fills the cylinders with pressurized fluid to prevent loss of discharge time due to fluid compression within each cylinder.

FIG. 11 schematically shows the configuration of the apparatus shown in FIG. 12, which allows various calculations and defines various values d, D, W, L, etc. for comparison. .

FIG. 12 is essentially a copy of FIG. 6 of German Patent Application Publication No. 3811625 to which the inventor is concerned, but it is a copy of FIG. If the leakage is too large and is a problem, the 12th
It is shown here to demonstrate that the illustrated configuration can be utilized.

However, based on the findings of the present invention, it is possible to add a second control section or an auxiliary control section to the configuration shown in FIG. It is possible to solve this problem. 14, 15, and 16 show configurations for forming auxiliary flows other than those described above, control units, or means for avoiding losses due to fluid compression. 12
This is a modification or alternative to the configuration shown in the figure.

FIG. 13 shows that the cylinder portions 14 and 1 are
A configuration example is shown in which control units 417 and 498 are provided to alternately send fluid to the 5.

FIG. 14 shows an example of a configuration in which two or more reciprocating motion control sections are provided according to the present invention.

One of these controls is for the main flow and the other for the auxiliary flow. It is also possible to provide only one reciprocating control for the main flow. FIG. 14 shows a rotating shaft 500 that is rotatably driven by, for example, the fluid motor or return flow motor 97 shown in FIG. 1 or 12, and the control member 41 is attached to this shaft.
7 and 617, or means for reciprocating either or both of them is provided. These control members 417, 617 are control area portions (radial enlarged diameter portions).
495, 496, 497, 498, 499, and these control areas allow fluid passages of 472, 408° 4
73, 474, 475, and 489'' respectively, or to keep these two fluid passages in communication at all times.

Note that in FIG. 13, two intermediate pressure cylinders are formed within a single cylinder block 418;
Further, the ends of the cylinder are closed by covers 512 and 513. This configuration, which consists of only a cylinder block and a cover, was proposed to the present inventor by Mr. Ozaki of RIKEN KIKI CO., LTD.

Figures 16 and 17 are XVIXV of Figure 14, respectively.
FIG. 2 is a cross-sectional view taken along the line I and XVII-XVII. 16 and 17, the main control member 417 in FIG. 14 is connected to the eccentric cam 501 via the connecting rod 503.
The second or auxiliary control member 617 is provided on the rotary shaft 500 and is driven reciprocatingly by the reciprocating speed of the sine function. A knock 502 located at the top indicates that the actuator is suddenly driven into axial movement and remains stationary during the rest of the same operating cycle. 16 and 12 further show detailed elements of the configuration, including a rotating shaft 500, an eccentric cam 50,
1, eccentricity e (521), pin 504, control member 41
7 end 505, connecting rod 503 with pin holes (eyes) 506, 507, outer surface 522 of the cam or recess 52
3. Contact surface 524.5 of end 508 of control member 617
25 etc. are included. In addition, Fig. 16 shows a recessed portion (undercut) for the purpose of creating a structure that suddenly connects the opening correctly in order to send a pressurized auxiliary flow to the cylinder portion.
524.526 can also be formed.

FIG. 15 shows that an overflow valve 550 can be provided between cylinders 14 and 15. Opening this valve 550 in a timely manner causes the still compressed fluid in one of the cylinders (cylinder 14 in FIG. 15) to expand into the other of them (cylinder 15 in FIG. 15); This other cylinder can be partially filled with fluid pressurized to a pressure between the desired high and low discharge pressures. Since it can be immediately filled with a fluid having a pressure approximately in the middle of , time losses due to fluid compression can be significantly reduced.

FIG. 15 also shows that the fluid pressurized to a pressure higher than the intermediate pressure is transferred to the pressure accumulator (
It is also shown that it can be configured to supply the accumulator) 553.

In this configuration example, a control unit 552 is provided, which allows excess pressure fluid to be supplied from the pressure accumulator 553 to the cylinders 14 and 1.
5 to quickly build up sufficient pressure in the cylinder to completely or at least partially eliminate the losses caused by compressing the liquid mentioned above.

Furthermore, FIG. 15 also shows structural features according to other important manifestations of the invention. At the moment of completion of the discharge stroke, some fluid still remains in the dead space and in a portion of each cylinder. When the control or control member begins to send return flow from the cylinder to the fluid motor 97 driven by the return flow, at that moment some fluid remaining in said cylinder and still under high pressure suddenly expands.

This is because the control or control member opens a fluid path to the fluid motor 97 which is driven by the low pressure return flow or return flow. Due to the rapid expansion of this fluid, the piston (
The piston assembly 5.8 or 6,9) is suddenly driven downwards. Then, due to the sudden downward stroke of the piston, a relatively large amount of fluid is rapidly supplied to the return flow motor (fluid motor) 97, and its movement (rotation) speed is accelerated. The resulting overspeed of the motor 97 causes the control device or control part 17 to rotate at an overspeed and prevent the opening of the fluid passage to the other cylinders at a premature point, i.e. before the dispensing process is yet completely finished. There is a risk that it may close. The inventors have discovered that as described above, the device may become ineffective and may be capable of dispensing only a fraction of the desired dispensing amount. By providing a second or auxiliary control as described above, such problems are reduced or eliminated. This auxiliary flow formation configuration has never been proposed before the present invention. The overflow valve (expansion fluid overflow valve) 550 shown in FIG. 15 also functions to reduce or prevent the risk of overspeed rotation of the return flow motor 97 as described above. This valve allows most of the inflation fluid to flow into one cylinder and prevents it from flowing into the return flow motor 97, which is then forced into a temporary overspeed force as described above. This is because full rotation is prevented.

In the configuration example shown in FIG. 16 in which the control unit 417 is driven by the eccentric cam 501, the movement of the control unit 417 is represented by the following equation: (14+, aω).

RX.

Sp (stroke) = R-Rcoscr+ -5ln”cr
Q31L R.

νp (velocity) = Rω (sin α +5tn2α)
(14) L Bp (acceleration) = Rω” (coscr + cos2c
r) (151 In the above formula, R = e in Figure 17, and L =
This is the distance between the axes of the connecting rod bottle holes 506 and 507.

FIG. 17 shows a calculated pressure curve for one operating cycle of a compressible fluid, such as oil or water, pump device as described above. This FIG. 17 shows 1000.2000.3000 and 40.0 when the eccentric cam drive of each control section or control member or reciprocating motion control section makes one rotation of the rotation angle α.
The respective curves are shown for a discharge pressure of 00 kg/c+ll. The dashed line ``expansion'' indicates the action of the expanding fluid.
"Compression" refers to the compression of the fluid that begins after the expansion fluid completes its expansion portion of the working cycle, and the solid line M-D
The actual pressures obtained when the overflow valve shown in FIG. 15 is provided and an overpressure auxiliary flow is provided to alternately fill cylinders 4 and 15 with fully pressurized fluid. discharge pressure).

Figures 18-A, 18-B, and 18-C show actual pressure (discharge pressure) graphs measured by an automatic recorder installed in the laboratory of RIKEN Instruments (1).

FIG. 18-A shows an overflow from the safety valve 409 shown in FIG. Appears in case of a configuration example in which it is not possible. Peak 561 (Figure 18-B) indicates overpressure at such a point in operation. Figure 18-B shows the pressure in the intermediate pressure chamber shown in Figure 1 at such time, and Figure 18-C shows the pressure in the discharge flow path beyond the high pressure outlet valve. Valley 56 of the curve in Figure 18-C showing pressure
2 shows the actual discharge losses caused by compression of the fluid if the inflation fluid overflow valve and auxiliary flow formation arrangement according to the invention were not provided. Figure 18-A, Figure 18-B
The figures and FIG. 18C clearly show that the phenomena and matters pointed out herein actually occur in the actual operation of high-pressure devices or high-pressure pump devices and in measurements related thereto.

For calculations regarding other points of technology that are relevant to the present invention, the following equation 06) for calculating the outflow rate from the pressure vessel body:
is also important.

Calendar Vout □ □ 14.14-rEji'HOeThis formula was derived by the inventor from the basic outflow calculation formula according to known technology.

FIG. 19 is a cross-sectional view of a common water pump that has been manufactured and sold for several decades and is an important part of the prior art. A crankshaft is provided within the housing 570 forming three eccentric bearings 571, 572 and 573 so that, via a connecting rod 574 and bearings 576, three drive pistons 575 are driven in reciprocating strokes. It has become. The piston 575 has a high pressure piston 57
A connection 577 to the high pressure piston 5 is provided.
78 is configured to reciprocate within the cylinder chamber or cylinder portion 579 of the high pressure cylinder 580. The cylinder head has an inlet valve 582 and an outlet valve 583.

In the suction stroke, water enters the cylinder chamber 579 from the supply pipe (line) 581 via the inlet valve 582, and in the discharge stroke, the large mouth valve is closed and the outlet valve is opened, and this water is discharged to the outside of the cylinder chamber. It will be done. This prior art water pump is well known to those skilled in the art.

FIG. 20 shows a calculation table (form) along with calculation formulas for calculating the piston stroke, the piston movement speed, and its acceleration.

FIG. 21 shows the same calculation table as shown in FIG. 20, with the calculated values actually calculated for the three-plunger type pump shown in FIG. 19 entered. This 21st
The table at the bottom of the figure also shows calculated values for a 9-plunger type pump, such as those shown in FIGS. 40 and 41.

Figure 22 shows the pistons 1.2 and 3 of the pump of Figure 19.
The discharge of is shown by curves 1.2 and 3.

What is not commonly recognized by users of the water pump shown in Figure 19, and perhaps ignored by the manufacturers of such equipment, is that the pump's delivery is highly uneven. It means that it is something. From the results of the exact calculations shown in curves 1.2 and 3 shown in FIG. 22 or in FIG. It can be seen that it exceeds 20%, albeit slightly.

Since the flow non-uniformity is large as mentioned above,
The pump of Figure 19 cannot be used for precision water jet cutting.

Moreover, such large flow non-uniformities have the disadvantage of requiring the use of very strong pipes on the discharge side, as normal pipes or hoses would be destroyed.

Under these circumstances, one of the objects of the present invention is to improve the flow uniformity of a plunge medium-sized pump device.

An example of an improved device according to the present invention in this regard is shown in the fourth example.
0 and 41. The pump device shown in these figures is a 9-plunger or 9-piston type pump device, the discharge characteristics of which are shown in FIG. The curves 9 to 9 in FIG. 24 indicate the discharge of the pistons 1 to 9 of the pump (device), and the curve shown at the top of the figure indicates that the crankshaft of the pump is rotated at the rotation angle α.
It shows the total discharge of all pistons when they rotate by .

The curve shown at the top of FIG. 12 shows a flow non-uniformity of about 2.03%. That is, the flow non-uniformity in the pump according to the present invention shown in FIGS. 40 and 41 is improved by about 10 times that of the conventional pump shown in FIG. 19. Therefore, the pump shown in FIGS. 40 and 41 can be effectively used for highly precise water jet cutting operations, and the effect of fluctuations in discharge flow on a pipe or hose is approximately 10 -(1/
10), making it possible for this pump to use ordinary pipes or hoses on its discharge side.

Now, Figure 23 shows the calculation formula for high pressure pipes or hoses.
A calculation table (form) and the results of calculations actually performed are shown in the table. Such calculations measure the strength of the high-pressure discharge line and also evaluate the non-uniformity (or non-uniformity) of the flow due to periodic expansion and contraction of the tube section between the axial ends of the line. necessary to do so. In this way, it is impossible to develop reliable high-voltage equipment without establishing a precise technical basis.

FIG. 25 shows a cross-sectional view of a conventional high-pressure booster device on the left side, and actual measurement data of high-pressure fluid discharged during one operation cycle of this device on the right side. In FIGS. 26 to 32 following FIG. 25, sectional views or schematic diagrams of example devices are shown on the left side, and high-pressure fluid discharged in one operation cycle is shown on the right side. Actual measurement data is shown.

FIG. 25 shows a conventional axial booster device with a fixed or constant drain or discharge pump 612 driven by a power source (plan) 611, which draws fluid from a tank 613 to It is sent to the controller (controller upper roller) 6I4, which controls the cylinders (sections) 603 and 60.
4, which drives the intermediate pressure pistons 605, 606 and thus also the high pressure pistons 607, 608. In this device, the pressure delivered during the working cycle is that indicated by P over time t, as shown in the right part of FIG. 25, with a trough upon compression of the fluid. It is a curved line. 2 shown below this pressure curve
Among the seed rectangles, the upper one represents loss due to fluid compression, and the lower one represents loss caused by fluid expansion (hereinafter referred to as loss).

The experimental results shown in FIG. 25 indicate that this conventional shaft booster device requires further improvement in order to eliminate or reduce losses and flow non-uniformity over time.

FIG. 26 shows how the device of FIG. 25 can be improved in accordance with one embodiment of the present invention.

In the embodiment shown in FIG. 26, according to the invention:
A one-way check valve 620 is provided in the suction flow path 619,
Further, fluid passages 617 and 618 are provided extending from the outlet opening of the valve 614 (control section 614 in FIG. 25) to the fluid passage above the check valve 620. The inflation fluid of each cylinder (section) is thus pumped to the pump PF and then to the inlet opening of the main control section. The inflation fluid from one of the two cylinders flows via a pump into the other cylinder, so that it is possible to fill this cylinder with a fluid which is precompressed to some extent. The result of this action is somewhat similar to that of the inflation fluid overflow valve 550 shown in FIG. The effect of this configuration according to the invention is that, as can be seen from the right part of Figure 26, the trough in the pressure over time curve never falls to the zero (0) value, and no losses due to liquid expansion are no longer observed ( (The rectangle shown at the bottom of FIG. 25 is completely eliminated), which means that the loss due to fluid compression is greater than in the case of FIG. 25, and is reduced to about 2. 25th
Note that the rectangle representing the force loss is smaller in FIG. 26 compared to the figure.

FIG. 27 shows yet another important embodiment of the present invention. In this embodiment, two power sources (plants) independently drive two pumps PF, PF. That is, power plants 632 and 633
drive pumps 630 and 631, respectively. Also, for the flow from each of these pumps, 2
independent controllers 634 and 635
Also provided. With such an independent configuration of the pump and control section, before the compression stroke of one piston (assembly) is completed, the control for sending pressurized fluid to the other piston (assembly) that will be operated next is performed. It becomes possible to start the operation of the unit. By the time the first piston reaches the end of its stroke, it can already begin to compress the fluid in the sill of the other piston. This can be understood by looking at the sloped line on the right side of FIG. However, the loss during compression of the fluid and the loss during expansion thereof have not yet been eliminated. This is clear from the rectangle shown on the right side of FIG. However, the discharge of high-pressure fluid is completely uniform with no fluctuations (troughs).

Therefore, the device in this case can be used for precise high-pressure water jet cutting with confidence. In the case of FIG. 27, there is a loss other than the above, namely, energy consumption when the second power source operates without being supplied with power. Although this loss is slight, it is illustrated by the horizontally elongated quadrilateral shown at the bottom of the right-hand portion of FIG.

The embodiment shown in FIG. 27 eliminates all flow non-uniformities and eliminates the need for expensive pressure accumulators, thus overcoming the major problems and disadvantages of prior art systems.

However, since the undesirable force loss (the loss indicated by the quadrilateral in the right-hand part of FIG. 27) still remains, this embodiment still leaves points to be improved.

The improvement required as described above is achieved by the second method according to the present invention.
This can be achieved by the configuration shown in FIG. According to this embodiment of the invention, the permanent sump pump (
Variable pumps Pv, ie, 640 and 641, are used in place of the pump 612) and control units 634 and 635 in FIG. These pumps are set to dispense at the moment of completion of the dispensing stroke of the previously actuated piston (assembly), just as the compression in the cylinder of the next actuated piston is just completed. In this embodiment as well, the second
As shown in the right part of FIG. 8, a completely flat characteristic line of the high pressure fluid flow without any fluctuations is obtained. Furthermore, losses due to fluid compression are reduced and force losses during fluid expansion are also reduced. These losses could be at least partially reversed because the inflation fluid flows into the pumps, allowing them to operate as fluid motors. In this way, the efficiency of the high pressure device was improved.

The example device shown in FIG. 29 is similar to that shown in FIG. 28, except that in the device of FIG. Sections 853 and 854 are provided. According to this configuration, during the expansion stroke, the pump is reversed and operates as a fluid motor, and expansion fluid from each cylinder flows into it, causing the pump to operate as a fluid motor, thereby providing power to the power source. It turns out. This eliminates the problem of loss when the power source operates under no load, and improves the efficiency of the entire device. Again, please refer to the right-hand portion of FIG. 29 where the quadrilaterals showing the discharge curves and force losses are shown.

In the example device shown in FIG. 30, instead of two power sources, only one power source (blanc) 611 is used,
No. 3,805, owned by the inventor hereby.
A unique IDEPU, such as the one disclosed in No. 675
It is configured to drive a pump. This pump, designated 644, is a reversible variable two-flow pump. Also in this device, a one-way check valve 620 is provided in the suction flow path 619 before reaching the branch flow path 646. Fluid flow from pump sections 644 and 645 is indicated by streams 647 and 648. According to this embodiment of the invention, losses in the idle power source are reduced and the inflation fluid is transported by one of the two pump parts operating as a fluid motor and the other of them operating as an intermediate pressure fluid delivery pump. At this time, it is possible to send it directly to the pump. As a result, losses during fluid compression can be reduced, or losses due to the power source can be similarly reduced. A highly efficient high pressure device with completely uniform flow can be realized.

FIG. 31 again shows an example of the conventional device. In this device, an intermediate pressure piston 650 is provided within a single intermediate pressure cylinder having two separate chambers 651 and 652 at each end of the piston. In this device, the expansion fluid acts directly as the driving fluid for the piston, and the two chambers
At the end of the piston stroke in one of the chambers 51 and 652, the compression of the fluid in the other chamber can be accelerated. There are still fluctuations in the discharge curve (right part of FIG. 31), but these are smaller than in the case of FIG. 25 and somewhat similar to the fluctuations in the case of FIG. 26. If this device incorporates an auxiliary flow formation arrangement according to the invention or an intermediate pressure accumulator, as described above, the non-uniformity of the flow can be reduced as indicated by the dashed line in the right-hand portion of FIG. You can.

FIG. 32 shows an example configuration that achieves the desired mitigation objectives as described above. This configuration is also in accordance with the invention, but in this case the power source (plant) is adapted to drive a variable reversible pump 653 and an overpressure filling pump H, designated by the reference numeral 658. At least one one-way check valve is also provided in this device in the suction channel before reaching the branch channel leading to the pump. This device is equipped with a cooler K, but if desired this can be provided in any device according to embodiments of the invention. Pressure accumulator 659
and 660 as an overpressure fill pump or auxiliary flow pump 6
It is connected to 58. When the variable reversible pump 653 is reversed at the end of a piston stroke in either direction, the accumulator 659 or 660 rapidly pumps fluid to the next cylinder to be actuated, causing the fluid to flow in that cylinder. Achieves rapid compression. This action takes place reversibly at the end of each piston stroke. As shown by the curves and quadrilaterals on the right side of FIG. 32, it is possible to achieve almost uniform fluid flow characteristics with very little fluctuation, and to reduce force loss to a negligible amount.

FIGS. 36 to 39 are overall views of pumps each showing a configuration in which the conventional pump shown in FIG. 19 is improved by the means of the present invention. Figure 36 shows RATI! A configuration incorporating the W device (system), FIG. 37 is an example of a configuration incorporating an ET opening device, FIG. 38 is an example of a configuration using an EPE device, and FIG. 39 is an example of a configuration incorporating a DBPEW device. An example of each configuration is shown. RATEW above, ET
EW, EPE- and DBPEI, each of which is an inventive device of the present invention, which is partially already disclosed in the patent application publication (announcement or publication).

FIG. 33 shows a configuration in which the arrangement of the two cylinders shown in FIG. 25 is reversed, in accordance with the present invention. According to this configuration, the large mouth valve 582,582 and/or the outlet valve 58
3.583 can be formed with a common inlet section 671 and outlet section 672.

In the configuration example shown in FIG. 34, two cylinders are arranged in parallel and are interconnected or communicated by an intermediate pressure fluid passage 673.

This configuration is also in accordance with an embodiment of the present invention.

FIG. 35 shows an inlet valve according to the present invention which makes it possible to minimize the space required by providing the inlet valve within the outlet valve.
Figure 3 shows the outlet valve configuration.

This valve arrangement may also be utilized in a variety of devices that utilize or handle compressible fluids such as air or other gases.

A particular advantage of this valve configuration is that dead spaces within the cylinder can be completely eliminated, thereby solving one of the above-mentioned main problems that the invention is aimed at. Inlet valve 681 and outlet valve 682 have a common intermediate axis. Since the piston head surface of the piston 606 moves very close to the head surface 683 of the inlet valve, the spacing between those surfaces is extremely limited, thus reducing the space 684 in the cylinder wall 601 to almost zero. ) can be done. The outlet valve 682 has a tapered seat 686 within the cylinder wall 601 while the large mouth valve 6
81 has a tapered seat 685 in the outlet valve that forms an angle of, for example, 90'' with respect to the tapered seat 686. Outflow flows from the opened outlet valve into the passage chamber 689 and from there to the outlet opening. 690. The incoming flow exits through the inlet opening 69
4 through the hollow interior of the outlet valve and then through the open artificial valve seat 685 into a chamber in the cylinder wall 601. Instead of the inlet opening 694, it is also possible to use an inlet opening 695 above the cover 696. A spring 689° and/or 692 can be provided for closing the response valve. Furthermore, the guide or insert member 691.687
It is also possible to provide a stop or retaining member 693, a valve stem 681'', etc. The valve arrangement described above can be partially arranged within the housing part 680.

In the pump shown in FIG. 36, a conical ring element 707, 708 of the RATEW device is provided between the drive piston 706 and the valve head with the mouth valve 702 and the outlet valve 703. Outlet valve 703 is fixed by retaining means 704 and 705. The reciprocating movement of the piston 706 causes the conical ring element to undergo compression and expansion (return from compression), which causes a periodic decrease and increase in the volume of the inner chamber 701, thus causing the inner chamber valve to Exhalation and suction of the fluid passing therethrough takes place. The RATEW device can be provided within a housing 700 for attachment to a pump or other device having a piston 706 as shown in FIG.

In the pump ETEW device shown in FIG. 37, one end of the inner chamber 701 of the conical ring element 708, 709 is supported on one end surface of the piston 711.

An outer chamber 710 is formed at the other end of the piston 711, and the piston 711 reciprocates in a liquid-tight manner to increase and decrease the volume of the outer chamber 710. A small diameter piston 706 is placed in an outer chamber 710.
The liquid in the outer chamber 710 acts on the large diameter piston 711. Compared to the piston RAT lung device of FIG. 36, the piston HTHW device of FIG. 37 allows for increased forces acting on the conical ring element.

The BPEW device in FIG. 38 uses a V-element 716 or other elements disclosed in patents owned by the inventor in place of the conical ring element described above. However, the element at one end of this EPE-device is a sealing bottom member 715, which seals off said one end of the inner chamber 701.

In the pump shown in FIG. 38 as well, the piston 706 moves into the outer chamber 714 to contract the (1) element and send out fluid from the inner chamber 701.

In the pump of FIG. 39, the DBPEW device is 2&[l
An outer chamber 717 is formed between the membranes or fluid separation members M, M. Fluid passage 718 and chamber portions 719,7
20 forms part of the outer chamber. Fluid passage 723
．． Parts 724 and 725 of the chamber 721 and 722 form two inner chambers which are connected to the large mouth valve 702.
and an outlet valve 703 (FIG. 36). The piston 706 moves into and out of the outer chamber 717;
In this configuration, the piston 706 extends deeply between the two inner chambers, creating a relatively long seal. Therefore, a relatively large amount can be discharged within a relatively limited space.

Fig. 40, along with Fig. 41, which is a cross-sectional view thereof, shows that, as previously shown in Fig. 24, the fluctuation in flow is very small at about 2%, and the flow of high-pressure fluid is highly uniform. An embodiment of the present invention is shown. Figures 40 and 41
In the figure, body parts 730, 731 and 732 are provided in axial front-to-back relationship, each of which has three working chambers 733, 734, 735; 736.
737, 738; 739° 740, 741, because the cylinder and piston are equally offset by 120°. Therefore, this device has 9 cylinders 733.734.735.736
．． 737.738°739, 740.741 are provided, among which piston 742.743.744
．． 745, 746, 747°, 748, and 749.750 perform reciprocating motion. These pistons are driven and guided over piston shoes 756, 756.756 by cylindrical but eccentrically arranged outer surfaces, also referred to as "travel guide surfaces" of eccentric cams 753.754.755. It has become. Eccentric cam 753.754.7
55 are attached to the solid shaft 751° at an angular interval of 120° in a front-to-back relationship in the axial direction. This axis 75
1° is the main axis of the device and is located at its center or center. Each cylinder can communicate with a chamber provided with a fluid separation member between the inner chamber and the outer chamber, as described above. In this case, chamber-forming housings 757, 758, 759 are formed as shown in FIG.
60 with bolts 761. Solid shaft 751
° is rotatably supported by a bearing 752.

A special effect of the configuration shown in FIGS. 40 and 41 is that
By one rotation of the solid shaft 751°, the next piston follows the piston and performs a rotational stroke of 40° of the rotation of this shaft 751°, and in this way the sine curve 1 shown in FIG. 24 is formed. As shown by ~9, the piston discharges fluid and the total discharge amount is like the uppermost curve in Figure 24, with the flow variation being slightly reduced to 2.03%. . In other words, with this device, approximately 98
% flow uniformity can be obtained, and the main objective of the present invention can be achieved. Additionally, by providing the cams and cylinders at angular intervals as described above, the balance of force can be made almost uniform in the radial direction, which has the effect of extending the useful life of the device. There is also.

FIG. 42 is a longitudinal sectional view showing yet another important embodiment of the present invention. The purpose of this embodiment is to solve two major problems in the prior art and thus achieve the two main objectives of the present invention. Valves in high-pressure pump systems are pressed against and isolated from their seats at high pressures or at high speeds and frequencies. For this reason, there is a limit to the useful life and reliability of the valve. Therefore, the purpose of this embodiment (FIG. 42) is to eliminate the need for a valve that is repeatedly pressed into and out of the seat portion as described above.

Another main objective is to eliminate the need for a reduction gear between the power plant and the pump and to obtain a compact high-pressure device with reduced weight.

In FIG. 42, the (rotation) shaft 763 is
62 bearing? 71 and 763. This shaft 763 is provided with a rotor portion 764 in which a hole or cylinder (portion) 769 is formed, in which a drive piston 768 reciprocates. A pivotable piston shoe 776 is provided on the front part of the piston 767, and the piston 767 is
The piston stroke can be configured to be guided directly on the inclined stroke surface 790 of the guide member 775 or indirectly through a piston shoe. Guide member 775
is within housing 762 and secured thereto. The piston shoe can further be configured to be guided by a rear retaining plate 780 at its rear shoulder. There are four balance parts (pocket l-) on the head surface of the piston shoe 776.
? 79 into which fluid communication 67? ? , 781
lubricating pressure fluid is pumped through 778 and 778. Fluid 1 passages 777, 781' and 778 are connected to the shaft 763,
Rotor portion 764 and/or drive member or drive piston 7
68. Although not specifically labeled and illustrated, a fluid passageway may be formed in a portion of each piston shoe 776. A fluid pressure adjustment (balancing) recess 781 may also be formed in the inner surface of the cylinder 769 of the drive piston or rotor portion 764.

Pressurized fluid can be sent to these recesses 781 from a branch flow path of the fluid passage 777 or 778.

A rear rotor section 765 is formed behind the front rotor section 764, which can also be referred to as a "high pressure rotor." This high-pressure rotor is provided with high-pressure cylinders (sections) 770 in the same number as the cylinders 769 and arranged at equal angular intervals behind them at the same angular intervals. A high pressure piston 7 is contained in the high pressure cylinder 770.
68 is provided so as to be able to reciprocate. High pressure rotor 76
A flow control section or control device 772 is provided at the rear of the flow control section 5, which is formed with a flow control inlet lower part 92 and an outlet part lower part 93. The rear end member or cover 777 of the housing 762 is formed with connecting lowers 86 and 787 for delivering fluid to and from the device and delivering high pressure fluid therefrom.

From what has been described above, the operating principle of the device shown in FIG. 42 will be rather simple and easily understood.

However, this device is based on the novel technical concept of the present invention, in which two rotors 764 and 765 are provided one behind the other in the axial direction, and a piston is provided in each of the cylinders.

The question with this device is whether it can operate not only as a low-pressure oil pump, but also as a high-pressure pump, even with non-lubricating fluids, such as water. In this latter case, the piston is likely to become welded to the cylinder, and so are the control surfaces to each other. Therefore, an invention is needed to prevent such welding or viscosity between adjacent parts or parts. The present invention provides a solution in this regard. FIG. 42 includes illustrations of special configurations according to the present invention as follows.

a) Providing two pistons: a drive piston 767 and a high pressure piston 768.

b) The drive piston is lubricated with lubricating oil, and both axial ends of the cylinder 769 in which this piston is installed are opened, thereby forming a hole that passes through the front rotor (section). .

C) The high pressure piston 768 and cylinder 770 have a smaller diameter than the drive piston, and the radial distance of the axes of these high pressure pistons and cylinders from the shaft and rotor is smaller than that for the drive piston and cylinder. To be there.

d) The drive piston does not pump fluid (intake and expulsion), the piston only acts to drive the high pressure piston into the discharge stroke.

e) A leakage fluid isolation member that forms leakage collection spaces 783 and 784 between the front rotor section 764 and the rear rotor section 765 and isolates lubricating oil leakage from high-pressure fluid leakage in these collection spaces. 782 was set up so that it could be obtained.

r) A positioning means, such as a thrust chamber 785, is provided to correctly position the shaft and rotor in the axial direction within the housing.

g) the flow control mechanism is prevented from rotating by means of restraints, and the flow control mechanism is provided with a thrust chamber in which the thrust body is located, thereby providing a problem-free seal between the stationary control surface and the rotating control surface; Special features such as providing fluid pressure and thrust means dimensioned and positioned to form a sealed configuration and prevent overthrust, tilting, etc. that could cause welding or sticking of the control surfaces together. is being configured.

Leakage fluid isolation or separation member 782 includes intermediate or end sealing members 791 and 7 for sealing one of the fluids from the other.
92, and this separation member 782 rotates with the rotor to send leakage radially outward by centrifugal force.
Each of the two types of leakage is sent to respective liquid collection chambers 783 and 784. The front head surface of the high pressure piston 768 abuts the rear end surface of the drive piston 767.

The flow control mechanism or control section 772 has a flow control inlet lower part 92 and an outlet part lower part 93 to feed fluid into and out of the high pressure cylinder 770.

It is possible to provide this control unit or control device 772 with a thrust chamber that is open at the rear and has a thrust body 773 inside, and the thrust body 773 allows fluid to flow into and out of the connecting lowers 86 and 787. The high pressure rotor 76 is sealed by sealing the spill and pushing the stationary control surface of the control section 772.
5. It can be configured to sealingly engage with the rotation control surface at the rear end.

With respect to the device shown in Figure 42, the sliding engagement may occur if the fluid is a non-lubricating fluid, e.g. water, or if the pressure is much higher than in a normal intermediate pressure fluid pump or motor. It is very difficult to prevent sticking or sticking or welding (or scratching) between control surfaces, so a great deal of "know-how" is required.

In the case of fluid equipment with a discharge pressure of up to 300 kg/cd,
Although the above-mentioned sealing arrangement between the control surfaces can solve the problem with the current state of the art, sealing is still not possible in the case of high pressures such as -barometric pressure or several thousand atmospheres. Also, if the fluid is a non-lubricating fluid, it is difficult to seal the control surface. When actually manufacturing the pump device shown in FIG.
I would like to be asked for some guidance.

FIG. 43 shows the predicted median values of compression percentage values for oil and water. This graph is important for understanding the present invention and is widely used in the inventor's laboratory.

However, the data shown in Figure 43 is not valid for all types of oil and water. Each time the type of water or oil changes, these data will be different, resulting in a curve that is more or less different from the curve shown in FIG.

FIG. 44 shows the results of calculations performed on the previously described portion of the present invention using the data in FIG. 43,
This figure shows one operating cycle, for example 36 cycles of an eccentric cam drive in the case of a rotary control device or control unit or a reciprocating control unit.
The pressure change during one revolution over 0° is shown. From this FIG. 44, the time loss during compression of the fluid within the chamber or cylinder can be understood.

FIGS. 45 and 46 are intended to provide a schematic explanation of what constitutes a pump device according to the present invention. These figures show a control device or control unit with a low-pressure opening and a high-pressure opening IP and that they are rotated by a motor. The piston is driven upward by fluid in the high pressure opening 11P and also pumps fluid downward from the low pressure opening during its downward stroke. Doil is the dead space (volume) filled with oil, and Dw
is the dead space (volume) filled with water. A membrane or other fluid isolation or separation member M separates water and oil from each other, and it is deflected within the working chamber. 45th
The figure shows the case where the member M is in the uppermost position, abuttingly supported on the upper stop surface, and below which the outer chamber is at its maximum volume. FIG. 46 shows the state in which the member M is in its lowest position and is abutted against the lower stop surface, thereby forming an inner chamber above it in the maximum volume state. . During each operating cycle of the device, member M undergoes a deflecting movement from the position of FIG. 45 to the position of FIG. 46 and vice versa. ■v and Ov indicate the population valve and outlet valve, respectively.

FIG. 47 is a schematic illustration of a BREW pump according to the present invention. Although this schematic diagram shows all the parts and parts shown on a single side of the drawing paper, in reality, some of the parts and parts are shown on the back side of the other spheres. It is in. Is the driving fluid, often intermediate or medium pressure oil, an intermediate pressure source? It is pumped by a pump P as IpS to a control body with a control opening or valve CV.

The fluid at intermediate pressure Mp is supplied through one control opening C during each actuation.
Through V, intermediate pressure cylinder? Flows into the IPC.

This fluid drives the intermediate pressure piston FIPK upwards, which in turn drives the high pressure piston also upwards or into the high pressure cylinder HPC. This high-pressure cylinder RPC is filled with low-pressure fluid via the intermediate chamber MC1 fluid passage and the large mouth valve B by a low-pressure fluid supply source (pump) tps when the high-pressure cylinder RPC is in its maximum capacity state. When the piston stroke begins, the large mouth valve B is closed and the fluid in the high pressure cylinder HPC is compressed to the high pressure HP. This high-pressure HP raw fluid is pumped into the outer chamber OC, and the membrane or member M
, thereby reducing the volume of the inner chamber IC and thereby forcing high pressure water (or other fluid) from the inner chamber IC through the outlet valve OV and into the discharge channel. At this time, population valve IV is closed.

When the above-mentioned discharge part (discharge stroke) of the working cycle is completed, the water supply (residential water pipe or low-pressure water pump) WpS
A big mouthful of water (or other fluids)! Send it towards V,
This valve is opened to send low-pressure LP fluid into the inner chamber IC, and the member M is pushed into the outer chamber OC to move the member M into the inner chamber IC.
Fill inside. Then, the oil in the outer chamber OC is forced into the high pressure cylinder RPC, thereby pushing the high pressure piston (IPK) downward. Fluid from the low pressure fluid supply pump LpS in the intermediate chamber MC promotes the downward movement of the high pressure piston HPK and the intermediate pressure piston MPK.

This forces the fluid in the intermediate pressure cylinder NPC through the other control opening or valve CV into the chamber of the return fluid motor D, which rotates the motor D and also rotates the control valve Cv communicating with this motor. . The used fluid flows out of the ERtV pump through the outlet opening. The code R is
This is a safety valve for preventing excessive pressure buildup in the intermediate chamber MC.

The apparatus shown in Figure 48 is substantially identical to the apparatus shown in Figure 12, except as follows.

12 is a patent application drawing showing the reference numerals and hatching, whereas FIG. 48 removes the reference numerals and hatching shown in FIG. It is an explanatory diagram showing a code. Further, in FIG. 48, for example, diameters d, D, etc. are further shown for reference in calculation. Figure 48 shows the complete ER of 1989-1990
Although the R-pump is shown, FIG. 47 only illustrates and explains the principle of operation of this device. It is also noted that FIG. 47 only shows one piston-cylinder set, whereas FIG. 48 shows both two piston-cylinder sets operating in alternating cycles. sea bream.

The foregoing portion of the specification has provided a portion of the basics of the invention and a detailed description of the technical invention, including several equations. In fact, the subject matter of the present invention is much broader and has been extensively researched. As they are too extensive to be fully described in this specification, only some of the results of calculations and research considerations will be presented, for example in the graphs such as Figures 49 to 53, Figures 44 and 66. Shown in the diagram.

FIG. 49 shows one operation cycle of the rotation control member, that is, 3
The pressure during rotation over 60° is shown. Discharge pressure is 100
The pressure changes are shown for each case of 0 atm, 2000 atm, 3000 atm, and 4000 atm.

FIG. 50 shows the change in discharge amount of the calculated, manufactured, and tested device example when fluid compression is taken into account.

FIG. 51 includes an illustration of the inflation fluid that actually occurs. The expansion fluid rotates the return fluid motor and prevents compression from immediately starting in the next cylinder to actuate. Compression begins after the fluid expansion of the first activated cylinder is completed, so the discharge is actually considerably less than that shown in FIG. 50. In Figure 51, approximately % of the discharge is lost. This means that for early 1988-1989 ERE-Pumps, 4
This means that only about twice the theoretical value could be discharged at 0,000 atmospheres. The results shown in Figure 51 are E
It shows that the RR-pump principle has been improved by the present invention. However, if the findings shown in FIGS. 49 to 53 had not been discovered at the stage leading up to the present invention, the present invention would not have been possible.

Fig. 52 shows the obtained pressures in one drawing, and Fig. 53 shows the pressures of 1000 atm, 2000 atm, 3
One working cycle is shown separately for each case of 000 and 4000 atmospheres.

FIG. 54 shows an enlarged view of the axial reciprocating control device or control section already illustrated. An intermediate pressure source (inlet opening) MpS is shown, and fluid passages Z1 and Z2 communicating with the first and second cylinders, respectively, are shown. The control member (controller) 801 includes control flanges 803 and 804 having control corners. The distance between the inner walls of the fluid passage 21 and Z2 is 806.
It becomes. The important point about FIG. 54 is that this distance 806 is substantially equal to the spacing or distance between the inner ends of the control angle of control member 801. - The middle part of FIG. 54 shows the state in which the control member 801 is in an intermediate neutral position, the upper part shows the left end position of this member, and the lower part shows its right end position.

55-58 depict a control housing similar to that shown in FIG. 54.

However, a control member 802 according to another embodiment of the present invention is incorporated within the housing shown in FIGS. 55-58. In this control member 802, the interval or distance between the inner ends of the control corners is 805, unlike 806 in the case of FIG. 54, and this distance [805! , the distance M80
It differs from the control member 801 in FIG. 54 in that it is larger than the distance between the inner wall surfaces of the fluid paths z1 and Z2. In FIG. 55, control member 8
02 is at its leftmost position, at this time fluid flows into the fluid passage z1 from the intermediate pressure inlet opening MpS and from the fluid passage z2 to the return fluid motor D (Fig. 47) or 97 (first
Figure). In FIG. 56, the control member 80
2 has moved slightly to the right. In this state, the control member 802 (in function, not design) corresponds to the inflation fluid overflow valve 550 shown in FIG. Z1
and the highly compressed fluid in the cylinder communicating therewith now begins to open and rapidly expand into the fluid passage z2;
The inside of the cylinder communicating with this passage Z2 is filled with compressed expansion fluid. Intermediate pressure inlet opening MpS (4th
A portion of the fluid from FIG. 7 also begins to flow into the fluid passage Z2.

In FIG. 57, control member 802 has moved further to the right. The overflow of inflation fluid from fluid passage z1 to 22 is almost complete. Fluid passage z1 begins to close and flow into the fluid passage leading to return fluid motor D or 97 (eg 301.302 in FIG. 1) is initiated. The flow from the intermediate pressure inlet opening MpS to the fluid passage Z2 and the cylinder communicating therewith continues. In Figure 58,
Control member 802 has reached its rightmost position. Medium pressure fluid flows from its inlet frontage MpS to fluid passage z2, while fluid from fluid passage z1 and the cylinder with which it communicates flows to return fluid motor D or 97. After this, the direction of movement of the control member or body 802 is reversed;
The second half or second stroke of the operating cycle is performed, and the reverse communication to the first half or first stroke described above is performed sequentially. In this manner, the configurations of FIGS. 55-58 provide a simple and reliable fluid overflow control mechanism or configuration according to the present invention.

Figures 59-64 illustrate one of the most highly developed "high tech" control mechanisms or devices according to the present invention. FIG. 44 is shown again above FIG. 46 to draw attention to the issue of time and pressure losses during each operating cycle. The purpose of the control devices shown in FIGS. 59-64 is to completely eliminate the above-mentioned losses, or to achieve a degree of perfection at least approaching that shown in FIG. 66.

The large loss area shown in FIG. 44 has been reduced to a small loss area between line E (expansion) and line A (compression or filling) in FIG. 66. This valuable result shown in FIG. 66 was achieved by the configuration shown in FIGS. 59 to 64, which will be described below.

The control device (controller) 810 is, for example, shown in FIG.
Rotation and reciprocation are driven by the configuration shown in FIGS. 63 and 64. A small pressure accumulator 811 is provided, which is filled by a pressure accumulation and filling pump 812. Furthermore, an internal chamber 813 is formed,
It includes a fluid passage 829 for receiving fluid from the fill pump 812, as well as a fluid passage 81 communicating with the controller.
4 and 815 are formed. A fluid passage 825 is provided between the filling pump 812 and the pressure accumulator 811,
The pressure accumulator also has a fluid passage 82 extending to the control device E810.
6 is also formed. At the rear end of the control device 810 is a guide slot 822, which will be described in more detail with respect to FIG.
and a guide finger member 820 which is rotatably supported by a bearing 821 and enters a guide slot 822 to cause the control device 810Z to pivot or rotate. Control device 8
10 has a number of specially and precisely positioned fluid passageways as shown by the dashed lines in FIG. In addition, in FIG. 59, the openings of these fluid passages are shown by solid lines.

In FIG. 60, a fluid passageway 816 passes through the control device 810 and communicates the interior chamber 813 with a fluid passageway that communicates with the cylinder Zl, and a fluid passageway 817 extends through the control bag W.
The cylinder Z2 is communicated with the pressure accumulation chamber 811 by penetrating through the cylinder 810. The pressure accumulation chamber 811 is connected to the return fluid motor D or 9.
7, the pressure accumulating chamber is filled with a filling pump 812 that can be driven indirectly by
Since it can be filled at a significantly higher pressure than S, for example twice that pressure, the fluid in this accumulator 810 will expand more rapidly into cylinder Z2 than fluid passage 817 at that higher pressure. Therefore, complete compression of the fluid within the cylinder occurs very quickly, so that losses due to fluid compression are significantly reduced, as shown in the left-hand portion of FIG. 66.

At the same time, cylinder z2, which is still in a compressed state,
Fluid within expands from fluid passageway 816 and then from controller 810 into interior chamber 813 . This allows for an expansion fluid pressure in the interior chamber 813 of the order of approximately % of the full compression pressure, thus forcing the filling pump 812 to a higher pressure compared to the low pressure, rather than at a pressure between the low pressure and the accumulator pressure. It can be operated efficiently at a pressure between the inflation fluid pressure and the pressure accumulator pressure. FIG. 61 is an explanatory diagram showing the direction of fluid flow by arrows and explaining that a certain distance in the axial direction, for example, a distance A is necessary between openings constituting alternate fluid passages; FIG. . FIG. 62 shows one operating state of the device shown in FIG. 60, which corresponds to the communication state during the second or latter half of the operating cycle, i.e., the pressure accumulator is connected to the cylinder by fluid passage 81B. The cylinder z2 is shown communicating with the internal chamber 813 through a fluid passage 819. Therefore, the communication state shown in FIG. 62 is the opposite of that shown in FIG. 60.

FIG. 63 shows the return fluid motor D as in the case of FIG.
97 indicates a configuration in which the (rotation) shaft 500 is rotated. Accordingly, the elements shown in FIG. 16 are designated by the same reference numerals in FIG. 63. However, FIG. 63 shows a more novel embodiment of the present invention.
It is different from the illustration.

First, the device in FIG. 63 includes a shaft 500 and a fluid motor 97.
Reduction gears 830, 831 and 832 are provided between them. This is important because even in the case of high-volume orbital motors, their rotational speed is too high for the ERE-pump. In the early ERE-pumps, the loss problems now discovered could not be adequately evaluated. The now discovered expansion fluid caused the control device to move too quickly, thereby preventing the piston, which performs the compression and displacement operations, from completing its stroke. This problem was overcome by providing the above speed reduction device. Second, the configuration shown in FIGS. 59-62 requires an accumulator/fill pump 812. Although this pump 812 is only a small pump, it produces approximately twice the pressure provided by the medium pressure working fluid source MpS. The pressure of the intermediate pressure source MpS is typically about 30
In this case, the discharge pressure of the filling pump 812 needs to be between 500 and 1000 atmospheres. This does not mean that pump 812 is expensive. Pump 812 is 1
．． - It continues to operate during the entire operating time of the operating cycle, and therefore the displacement per revolution is rather small. This means that although the force of this pump 812 may be small and the delivery pressure may be high, it may deliver only a small amount of fluid. In the configuration of FIG. 63, the pump 812 reduction gear f
It is designed to be driven by a motor 97 via 830 and 831. Third, since the controller 810 needs to be capable of not only reciprocating motion, but also rotation in order to enable precise fluid passage communication at precise times, the controller 810 and Drive means 833-
834-500 (these are means similar to those in FIG. 16), the reciprocating drive train 500-833
- It is necessary to provide means for enabling the control device 810 to rotate relative to the control device 834. Therefore, the control device 810 is provided with a flange 837 at one end, which is attached to the cover 840.
It is supported by bearings 836 and 838 within a reciprocating housing 839 having a diameter. With this configuration, the control device 810 can reciprocate together with the housing 839 and can rotate relative to the non-rotatable housing 839.

FIG. 64 is basically an enlarged partial view of the left end portion of FIG. 59 viewed from above. This FIG. 64 shows the guide finger member 8.
20 is shown retained within the guide slot 822.
This figure also shows that a portion of the guide slot 822 is
10 is also shown to be inclined with respect to the axial direction of the reciprocating motion. Positions 841 and 842 indicate the points at which accumulator 811 communicates with cylinders Zl and Z2. 180
and 360 indicate the degree of pivoting, which in this case is not a pivoting motion but a complete rotational motion. Complete reciprocating motion 1
In a cycle, controller 810 makes one complete revolution of guide slot 822, shown in FIG. However, this movement may also be pivoting if necessary. At the bottom of FIG.
It is the rightmost position of No. 22.

From this state, the control device moves to the right. Since the corner between the bottom portions of the guide slot is below the central axis of the guide finger, the guide finger does not enter the downwardly sloped slot portion of the bottom, but rather enters the upwardly sloped slot portion and When the control device moves to the right, this results in a relative movement of the guide finger 820 to the left with respect to the guide slot 822. When finger member 820 moves to the left within slot 822, it pushes the slot downward;
This causes the control member 810 to rotate.

In the center of FIG. 64, the finger member 820 is at the leftmost position, and from this state the direction of reciprocating motion of the control device 810 is reversed. Finger member 820 then fits into slot 822
Enter the upper part and continue to rotate the control until it reaches the upper position of Figure 64, and at the point of 360° rotation, the
The state is similar to that at the time of 0° rotation of the bottom portion in FIG. 4. With this configuration, the communication states shown in FIGS. 55 to 62 are realized. Note that if the control device is configured to perform only reciprocating motion, it cannot be activated.If the control device is capable of performing only 0-rotational motion, it can be activated, but in this case, a complicated A layout configuration is required. This is because it is necessary to communicate the accumulator to different cylinders at different parts of a complete cycle of operation.

FIG. 65 shows that in a further embodiment of the present invention, high pressure pistons 5 and 6 may be disposed facing each other between two working chambers, each consisting of an inner chamber IC and an outer chamber OC. It shows what is possible. cylinders 11 and 1
2 communicates with the two working chambers through fluid passages 846 in mutually opposite directions. Valve 68 shown in FIG.
1.682 and 683 can be provided in each inner chamber IC.

67 and 68, the bottoms of the high-pressure devices or pumps shown in FIGS. 1 and 12 are removed and these devices or pumps are equipped with the controls shown in FIGS. 26-32. Alternatively, it has been shown that it is possible to selectively incorporate control means and pump devices. In this case, the sixth step is to seal the intermediate pressure cylinders 14 and 15 with the cover 873.
7 and 68 show. Fluid passages 868 and 869 for intermediate pressure fluid are provided in this cover 873, and these are connected to the pumps shown in FIGS. 26, 27, 28, 29, 30, 31, and 32. or in communication with the control section of the control mechanism or pump PCMP or 870.871.

Figures 67 and 68 also show that in this configuration the intermediate pressure pistons 8 and 9 are preferably provided with piston rods 860, 861 extending rearwardly or out of the device. Attach adjustable position confirmation members (markers) 862, 863 to these piston rods at fixed positions in their axial direction. For example, holders (holders) 866, 867
Sensors 864 and 865 are fixed to the support so as to be adjustable in position in the axial direction of the support. 67 and 68 on either the position confirmation member 862, 863 or the sensor 864, 865.
Means are provided for automatically causing axial displacement according to the pressure within the device as shown in the figures. According to the present invention, before a fully-actuated piston completes its dispensing stroke, a command is sent to the next pump or controller, and when a fully-operated piston completes its dispensing stroke, the next The actuating piston has already entered its cylinder sufficiently deep to obtain sufficient pressure for discharge, as shown in the right-hand portion of either of Figures 26-32. The position confirmation member 86
2.863 and sensors 864 and 865 are placed and fixed.

The sensors 864,865 may be, for example, light beams, thereby causing electronic or electrical devices or
The structure may be configured to control the force adjustment position of the variable pump shown in FIG. 2, or the structure may be configured to energize a solenoid for operating the control unit or control device shown in FIGS. 26 to 32. is possible. For activation of the return stroke of the piston which is to be operated next, a sensor and/or a position confirmation element as described above may be further provided, and this sensor and/or position confirmation element may be used as a timing device, e.g. a time switch. It is also possible to replace it with

FIG. 69 shows an alternative configuration of the intermediate portion of main housing 464 shown in FIG. In the configuration example shown in FIG. 69, in addition to the high pressure piston 5.6, the piston means 555
5.6666 are provided.

The diameters of these piston means and the diameter of the high pressure piston 5.6 are different. As in the case of the configuration example shown in FIG. Since the diameter of the seal differs from that of the piston 5.6 and therefore the diameter of the seal likewise differs, the seal in this example has the reference numeral 1453.1.
454.2453 and 2454. The stroke space or piston chamber 1455, 1456 has a length of 872
, which is preferably the same length as the piston stroke distance. These stroke spaces 1455, 1456 are designed such that the parts of the piston that are affected by two different fluids are never equal. These stroke spaces communicate with low pressure or leakage collection chambers by liquid passages 1457.

The apparatus shown in FIG. 70 has an outer chamber 874 formed at the forward or discharge end of the high pressure piston 5.6. Inside these chambers 874, there are provided a large number of axially flexible conical rings L-878, 879, which are formed in pairs with their radial ends sealed together. . The above-mentioned end portions may be sealed together by plasma welding, for example. For the conical annular element or accordion-shaped fluid separation member 878,879, the outer chamber 8
74 is sealed by a bottom plate 877, and the other end of the fluid separation member 878,879 is provided with a radially widened conical ring whose radially outer edge 880 connects the housing 464 and the inlet. It is sandwiched between the valve 38 and the outlet valve 39 and is sealed and fixed. The sealed inner chamber 875 of the accordion-shaped fluid separation member 878,879
It is. In order to avoid dead space that would cause a reduction in discharge volume and efficiency, a portion of the inner chamber 875 is occupied by a filler 896, which includes the inlet valve 38 and the filler 896. A fluid passageway 897 can be provided. In FIG. 70, the piston 6.9 on the right side has completed its discharge stroke. In this state, the accordion-shaped fluid separation member is in a completely compressed state, and the bottom plate 877 is in close proximity to the bottom surface of the filling member 896. In the left-hand part of FIG. 70, the pistons 8, 5 are in a position where they have completed their return stroke. The adion type fluid separation member is in a fully expanded or inflated state. The effect of the configuration shown in FIG. 70 is that the piston stroke path and discharge amount can be increased compared to the configuration example shown in FIG. 12.

The degree of this effect is determined by the conical ring elements 878, 8
79, but in the case of a high pressure device as shown in FIG. 70, it is necessary to select appropriate materials and perform appropriate welding or brazing.

In the configuration example shown in FIG. 71, the membrane member 5 is attached to the side or end.
An intermediate plate member 881 having a diameter of 8.58' is connected to the housing 46.
4 and 464'. housing 4
64.464' are formed with high pressure cylinders 11, 12 and are connected to an intermediate pressure source and a control or control device as described above. In Figure 71,
Housings 464 and 46 provided axially opposite each other
Although only a small portion of the 4'' is shown, these /S housings and the kind of components provided therein or on them are shown in the drawings and already described above to which reference has been made. did. Housing 464°46
4'', membrane member 58, 58', and intermediate housing are integrally assembled with bolts, but illustration of these bolts is omitted in FIG. 71. Intermediate plate member 88
Fluid passages 882, 883 are formed through 1 and are connected to a port valve 38, 38' and an outlet valve 39, 39'.
are provided for each.

The valve housings 892, 893 can each have an inlet fluid passageway 894 and an outlet fluid passageway 895. The outer chambers 884 and 885 each have a cylinder (section) 11
and 12. Because of this communication configuration, the hole (bore)
888.889 is formed. Inner chambers 886 and 887 communicate with fluid passages 882 and 883, with fluid passage 882 communicating with one inner chamber 88.
6 , while the other inner chamber 887 communicates with fluid passage 8
83.

For these communication configurations, multiple, very small diameter holes (
A bore 890 is formed from the inner chamber 886 to the fluid passageway 882, and a number of similar holes 891 are formed from the inner chamber 887 to the fluid passageway 883. For 4000 atm water pump equipment, the above 6890
and the diameter of 891 must not exceed 0.8 mm,
The diameter of 6888 and 889 may be 2 to 4 mm.

In the configuration example of FIG. 71, the illustrated membrane member 58°58°
Instead, it is possible to use an accordion-type fluid separation member or a conical ring element 878, 879 shown in FIG. 70. This fluid separation element can also be used in the embodiments shown in the other figures already referred to, which are suitable in principle for providing this accordion-type fluid separation element.

It is understood that the claims at the beginning of this specification contain detailed descriptions of the invention other than those described above, and that such descriptions also constitute a part of the description of the preferred embodiments of the invention. I hope that.

Furthermore, in addition to what is described in the claims, the present invention has various features described in the following additional claims.

Supplementary Note 1: The device according to claim 1, wherein the auxiliary means is an expansion fluid overflow valve that sends a portion of the fluid of one of the two pairs of cylinders to the intermediate pressure cylinder of the other pair of cylinders. A high-pressure device characterized by:

2. The high-pressure device according to appendix l, characterized in that the auxiliary means is an auxiliary flow forming means that temporarily fills one of the intermediate pressure means.

3. The apparatus of claim 1, wherein the auxiliary means comprises one pump for each of the two pairs of cylinders, one of the two pumps delivering fluid to the intermediate pressure cylinder with which it still communicates. A high-pressure device characterized in that it is formed by providing control means for sending fluid from the other pump to the cylinder with which it communicates while the other pump is in communication.

4. The device according to claim I, wherein a pressure response control device is provided in combination with the two pairs of cylinders, and the piston is controlled and driven so as to alternately perform a discharge stroke and a return stroke, and the piston one of which forms a first piston and the other of which forms a second piston, respectively;
the second piston following the first piston and the first piston following the second piston; one of the control means causing the second piston to move in accordance with the first piston stroke; Start the discharge stroke of the piston,
Further, the other of the means is configured to start the discharge stroke of the first piston in accordance with the stroke of the second piston, and the first piston, the second piston, and the control means are arranged so that the first piston, the second piston, and the control means are configured to start the discharge stroke of the first piston in accordance with the stroke of the second piston. The dispensing stroke of the next piston is started during part of the stroke of the next piston, so that at the moment the first piston completes its dispensing stroke, the fluid in the cylinder of the next piston is already sufficient or 1. A high-pressure device, characterized in that it is provided with pressure control means for achieving a completely compressed state.

5. The device according to Supplementary Note 4, wherein a position confirmation member (marker) is provided on the intermediate pressure piston, and in response to this member, the piston that performs the next stroke movement starts the discharge stroke. A high-pressure device characterized by being equipped with a sensor that issues commands.

6. The device according to claim 4, further comprising a control unit or a control device in which the movable part is driven by a fluid motor driven by the return flow from the cylinder, so that the fluid is fed by the control unit or control device. A high-pressure device characterized by:

7. The high-pressure device according to claim 4, characterized in that sealing means are provided between the leakage collection chambers.

8. The apparatus according to claim 4, wherein a piston stroke response sensor energizes the flow direction control means for controlling the flow of pressurized fluid alternately and periodically to the intermediate pressure cylinders of the two pairs of cylinders. A high-pressure device characterized by being provided with.

9. The device according to claim 4, characterized in that the piston that performs the thrust operation later is communicated with the intermediate pressure fluid before the piston that performs the thrust operation first completes its pressure stroke. High pressure equipment.

10. The apparatus according to claim 5, wherein a piston stroke response sensor energizes the flow direction control means for controlling the flow of pressurized fluid alternately and periodically to the intermediate pressure cylinders of the two pairs of cylinders. A high-pressure device characterized by being provided with.

11. Apparatus as claimed in claim 5, including a second control or control device for directing an auxiliary flow to the intermediate pressure cylinder when the piston is in its rearmost position, thereby rapidly moving the intermediate pressure piston. A high-pressure device characterized in that the high-pressure piston associated with the high-pressure piston is driven by thrust to advance deeply into the cylinder, thereby pre-compressing the fluid in the high-pressure cylinder, thereby speeding up the compression of the fluid in the high-pressure cylinder. .

12. The high-speed device according to claim 5, wherein the control means comprises a reciprocating motion control member.

13. The device according to claim 5, characterized in that the piston that performs the thrust operation later is communicated with the intermediate pressure fluid before the piston that performs the thrust operation first completes its pressure stroke. High pressure equipment.

14. The device according to claim 5, characterized in that the pistons are provided oppositely on a common axis, and that an intermediate plate member is provided between the two housings containing these pistons. Takasho equipment.

15. The device according to Supplementary Note 14, wherein two independent fluid passages are formed in the intermediate plate member, each having a diameter of 2 m extending from the inner chamber to the fluid passages adjacent to each of them.
A high-pressure device characterized in that a hole of less than m in size is formed.

[Brief explanation of the drawing]

Figure 1, Figure 2, Figure 3, Figure 4, Figure 5, Figure 6, Figure 7, Figure 8, Figure 9, Figure 12, Figure 13, Figure 14, Figure 15. 16, 17, 67, 68, 69, 70 and 71 are cross-sectional views or partial cross-sectional views of the high-pressure device according to the present invention. 10 and 11 are cross-sectional or partial cross-sectional views of prior art devices. FIGS. 15 and 16 are partial cross-sectional views showing partial structural embodiments of the present invention. Fig. 17, Fig. 18-, Fig. 18-B, and Fig. 18-C are diagrams (graphs) for explaining the present invention in detail, and diagrams (graphs) showing measurement results by an automatic recorder. graph). 19th
The figure is a sectional view showing a conventional water pump. Figure 20,
Figure 21, Figure 22, Figure 23, Figure 24, Figure 25,
Figure 26, Figure 27, Figure 28, Figure 29, Figure 30,
Figure 31, Figure 32, Figure 43, Figure 44, Figure 45,
Fig. 46, Fig. 47, Fig. 48, Fig. 49, Fig. 50,
Fig. 51, Fig. 52, Fig. 53, Fig. 54, Fig. 55,
FIG. 56, FIG. 57, FIG. 58, and FIG. 66 are explanatory diagrams for understanding the basic technical contents included in the present invention. Figure 33, Figure 34, Figure 35, Figure 36,
37, 38, 39, 40, 41, and 42 are partial cross-sectional views showing other partial structural embodiments of the present invention. FIGS. 59, 60, 61, 62, 63, 64, and 65 are partial cross-sectional views showing still other partial structural embodiments of the present invention. II, 12.14.15.733.734.735.
736.737°738, 739.740.741...
- Cylinder or cylinder part (11, 12...high pressure cylinder, 14.15...intermediate pressure cylinder), 5.
6. 8. 9.742.743.744°745,
746.747.748.749.750... Piston (5°6... High pressure piston, 8.9... Intermediate pressure piston), 17, 772... Control unit, control device or control mechanism, 517 ... Auxiliary control unit, 753.754.
755...Eccentric cam, 751'. 763... shaft (751'... solid shaft, 763...
Rotating shaft), 464762...Housing, 764...
・Fluid handling rotor or simply rotor, 765...Rotor with built-in drive piston or simply rotor part, 775...(
piston stroke) guide member, 783.784... (leakage) collection space, chamber or groove. Book l→↓shi l6Q, 560- Fig, BaFt″9,21 1wrns (ksu1ρ-po) 3 places/claw Puguρ FrtatnrdrJtIlgel, sr;z4eR Ft, wings Flν, 52 Ft'1.53

Claims (1)

[Claims] 1. A high-pressure device comprising two pairs of cylinders in which pistons that perform reciprocating motion are fitted in a sealed manner, and a control means that alternately sends fluid under pressure to parts of these cylinders. and each pair of cylinders is comprised of an intermediate pressure cylinder having a first diameter and a high pressure cylinder having a second diameter smaller than the first diameter, and each pair of cylinders is comprised of an intermediate pressure cylinder having a first diameter and a high pressure cylinder having a second diameter smaller than the first diameter; consisting of an intermediate pressure piston and a high pressure piston having a diameter corresponding to the diameter and the second diameter; the fluid in the cylinder is a liquid; and the liquid in the high pressure cylinder is temporarily pressurized to a pressure exceeding 500 atmospheres. and auxiliary means for fluid compression are provided, which compress the fluid in the high-pressure cylinder at a higher rate than the average flow rate of the intermediate-pressure fluid delivered to the intermediate-pressure cylinder, so as to reduce the fluid to an outlet pressure of the high-pressure cylinder. A high-pressure device characterized by being configured to compress to 2. A high-pressure device comprising a shaft rotatably supported in the center of the device, and a cam having an eccentric guide surface for driving a piston that reciprocates within a cylinder is attached to this shaft. , the cams are composed of three cams arranged in the axial direction with an angular interval of 120° from each other, and the cylinders are composed of three cylinders in the first group and three cylinders in the second group. and a third group of three cylinders,
three cylinders of each group are formed at an angular spacing of 120° with respect to each other, and their axes are arranged on an imaginary plane perpendicular to the axes of said axes; imaginary surfaces of the cylinders are axially spaced apart from each other, each group of cylinders is provided with one such cam, a part of which is placed on the imaginary surface of the cylinder group; A high-pressure device characterized in that the pistons therein are sequentially driven at an angular difference of 40° with respect to the rotation of the shaft. 3. A fluid handling rotor rotatably supported in the housing, a piston stroke guide member provided in the housing, an inlet and an outlet opening formed in the housing, which sucks in and discharges fluid, and performs reciprocating motion in the cylinder. A high-pressure device having a control means for controlling a discharge piston and a fluid flow, wherein a rotor with a built-in drive piston is provided in the axial direction of a fluid handling rotor, and the rotor with a built-in drive piston has at least indirectly the above-mentioned A drive piston is provided axially projecting from the bore to engage the piston stroke guide member and form a rear end surface; a hole in the rotor incorporating the drive piston is provided at a first distance from the axis of the fluid handling rotor, and a cylinder is provided at a second distance from the axis of the fluid handling rotor; A high-pressure device characterized in that the distance is larger than the second distance. 4. A high-pressure device comprising two pairs of cylinders in which reciprocating pistons are fitted in a sealed manner, and a control means for alternately supplying fluid under pressure to parts of these cylinders, the apparatus comprising: a pair consisting of an intermediate pressure cylinder having a first diameter and a high pressure cylinder having a second diameter smaller than the first diameter; the pistons each having the first diameter and the second diameter; an intermediate pressure piston and a high pressure piston having diameters corresponding to A high-pressure device characterized by the following: 5. A high-pressure device comprising two pairs of cylinders in which reciprocating pistons are fitted in a sealed manner, and control means for alternately supplying fluid under pressure to parts of these cylinders, the apparatus comprising: a pair consisting of an intermediate pressure cylinder having a first diameter and a high pressure cylinder having a second diameter smaller than the first diameter; the pistons each having the first diameter and the second diameter; an intermediate pressure piston and a high pressure piston having diameters corresponding to A high-pressure device characterized in that a flexible chamber separation or fluid separation member such as a membrane member is provided between two chambers.